PARAMETRIC DESIGN 11.1 NOMENCLATURE
TABLE 11.VII STRUCTURAL WEIGHT COEFFICIENT K (1, 18)
11.4 HYDRODYNAMIC PERFORMANCE ESTIMATION
right in Figure 11.18. The resulting weights and centers are linked directly to the italicized weights and centers entries in the WEIGHTS I spreadsheet summary. Inputs needed for these design models are entered on the linked Weights and Centers Estimation spreadsheet.
11.4 HYDRODYNAMIC PERFORMANCE ESTIMATION
The conceptual design of a vessel must utilize physics-based methods to simulate the propulsion, maneuvering, and seakeeping hydrodynamic performance of the evolving design based only upon the dimensions, parameters, and intended features of the design. An early estimate of resistance is needed in order to establish the machinery and engine room size and weight, which will directly influence the required overall size of the vessel. Maneuvering and seakeeping should also be checked at this stage of many designs since the evolving hull dimensions and parameters will affect this performance and, thus, the maneuvering and seakeeping requirements may influence their selection. This Section will illustrate this approach through public domain teaching and design software that can be used to carry out these tasks for displacement hulls. This available Windows software environment is documented in Parsons et al (40). This documentation and the compiled software are available for download at the following URL:
www-personal.engin.umich.edu/~parsons 11.4.1 Propulsion Performance Estimation 11.4.1.1 Power and Efficiency Definitions
The determination of the required propulsion power and engine sizing requires working from a hull total tow rope resistance prediction to the required installed prime mover brake power. It is important to briefly review the definitions used in this work (41). The approach used today has evolved from the tradition of initially testing a hull or a series of hulls without a propeller, testing an individual or series of propellers without a hull, and then linking the two
together through the definition of hull-propeller interaction factors. The various powers and efficiencies of interest are shown schematically in Figure 11.19. The hull without a propeller behind it will have a total resistance RT at a speed V that can be expressed as the effective power PE,
PE = RT V/ 1000 (kW) [57] where the resistance is in Newtons and the speed is in m/s. The open water test of a propeller without a hull in front of it will produce a thrust T at a speed VA with an open water propeller efficiency ηo and this can be expressed as the thrust power PT,
PT = TVA / 1000 (kW) [58] These results for the hull without the propeller and for the propeller without the hull can be linked together by the definition of the hull-propeller interaction factors defined in the following:
VA= V(1 – w) [59] T = RT/ (1 – t) [60] ηP = ηoηr [61] where w is the Taylor wake fraction, t is the thrust deduction fraction, ηP is the behind the hull condition propeller efficiency, and ηr is the relative rotative efficiency that adjusts the propeller’s open water efficiency to its efficiency behind the hull. Note that ηr is not a true thermodynamic efficiency and may assume values greater than one.
Substituting equations. 59 and 60 into equation 58 and using equation 57 yields the relationship between the thrust power and the effective power,
Figure 11.19 - Location of Various Power Definitions
PT = PE (1 – w)/(1 – t) [62] from which we define the convenient grouping of terms called the hull efficiency ηh,
ηh = (1 – t)/(1 – w) = PE/PT [63]
The hull efficiency can be viewed as the ratio of the work done on the hull PE to the work done by the propeller PT. Note also that ηh is not a true thermodynamic efficiency and may assume values greater than one.
The input power delivered to the propeller PD is related to the output thrust power from the propeller
PT by the behind the hull efficiency equation 61 giving
when we also use equation 63,
PD = PT /ηP = PT /(ηoηr) = PE /(ηhηoηr) [64] The shaft power PS is defined at the output of the reduction gear or transmission process, if installed, and the brake power PB is defined at the output flange of the prime mover.
When steam machinery is purchased, the vendor typically provides the high pressure and low-pressure turbines and the reduction gear as a combined package so steam plant design typically estimates and specifies the shaft power PS, since this is what steam turbine the steam turbine vendor must provide. When diesel or gas turbine prime movers are used, the gear is usually provided separately so the design typically estimates and specifies the brake power PB, since this is what prime mover the prime mover vendor must provide. The shaft power PS is related to the delivered power PD transmitted to the propeller by the sterntube bearing and seal efficiency ηs and the line shaft bearing efficiency ηb by,
PS = PD/(ηsηb) [65] The shaft power PS is related to the required brake power PB by the transmission efficiency of the reduction gear or electrical transmission process ηt by,
PB = PS/ηt [66] Combining equations. 64, 65, and 66 now yields the needed relationship between the effective power PE and the brake power at the prime mover PB,
PB = PE /(ηhηoηrηsηbηt) [67]
11.4.1.2 Power Margins
In propulsion system design, the design point for the equilibrium between the prime mover and the propulsor is usually the initial sea trials condition with a new vessel, clean hull, calm wind and waves, and deep water. The resistance is estimated for this ideal trials condition. A power design margin MD is included within or applied to the predicted resistance or effective power in recognition that the estimate is being made with approximate methods based upon an early, incomplete definition of the design. This is highly recommended since most designs today must meet the specified trials speed under the force of a contractual penalty clause. It is also necessary to include a power service margin MS to provide the added power needed in service to overcome the added resistance from hull fouling, waves, wind, shallow water effects, etc. When these two margins are incorporated, equation 67 for the trials design point (=) becomes,
PB(1 – MS) = PE (1 + MD)/(ηhηoηrηsηbηt) [68] The propeller is designed to achieve this equilibrium point on the initial sea trials, as shown in Figure 11.20. The design match point provides equilibrium between the engine curve: the prime mover at (1 – MS) throttle and full rpm (the left side of the equality in equation 68), and the propeller load with (1 + MD) included in the prediction (the right side of the equality).
The brake power PB in equation 68 now represents the minimum brake power required from the prime mover. The engine(s) can, thus, be selected by choosing an engine(s) with a total Maximum Continuous Rating (or selected reduced engine rating for the application) which exceeds this required value, MCR ≥PB = PE(1 + MD)/(ηhηoηrηsηbηt(1 – MS)) [69]
Commercial ship designs have power design margin of 3 to 5% depending upon the risk involved in not achieving the specified trials speed. With explicit estimation of the air drag of the vessel, a power design margin of 3% might be justified for a fairly conventional hull form using the best parametric resistance prediction methods available today. The power design margin for Navy vessels usually needs to be larger due to the relatively larger (up to 25% compared with 3-8%) and harder to estimate appendage drag on these vessels. The U. S. Navy power design margin policy (42) includes a series of categories through which the margin decreases as the design becomes better defined and better methods are used to estimate the required power as shown in Table 11.IX.
TABLE 11.IX - U.S. NAVY POWER DESIGN